Multi-stage radial flow turbine

ABSTRACT

Various multi-stage radial turbine configurations that provide highly efficient momentum transfer between a fluid and the mechanical interface in both power producing and power consuming undertakings.

The present invention is a continuation-in-part of U.S. patentapplication Ser. No. 14/209,019 filed Mar. 13, 2014, which in turnclaims priority on U.S. Provisional Patent Application Ser. No.61/782,044 filed Mar. 14, 2013, which are all incorporated herein byreference.

The present invention is directed to a multi-stage radial turbine forusage in energy capture from fluid streams with low to moderate specificspeed. The turbine can also be driven and used as a very efficientliquid pump and gaseous phase fan or compressor. Additionally, combiningcompression, a burner, and energy capture creates a new type of internalcombustion engine. FIG. 1 introduces and illustrates most simply some ofthe wide range of configurations that can utilize this new type ofturbine.

BACKGROUND OF THE INVENTION

Various radial designs for a turbine have been utilized since the dawnof the industrial revolution and their characteristic flow is found innumerous patents. FIG. 2 provides a qualitative graph that is slightlymodified from that found in “Mechanics of Fluids”; author: IrvingShames; ISBN 0-07-056385-3, page 616.

The graph illustrated in FIG. 2 shows through curves of efficiencyversus specific speed whether the optimal flow regime is either radial,axial or the transitional range between these two. At low speeds, theradial orientation provides superior efficiency. This graph is aqualitative selection tool. The calculations required to quantify thespecific speed can be rather complex and include the fluid flow rate andphysical properties as well as the mechanical geometry and rotationalspeed. Without usage of any quantitative calculations, deductivereasoning can be used to categorize wind as having a very low specificspeed. In the case of wind energy, a proposed site's wind speeddistribution is important data for any type of economic feasibilitystudy. Average wind speeds can be mapped into regions and givenclassification values that typically fall from one to five. Axial flowwind turbines are typically built in class four regions. Class threeregions are considered a minimum requirement for these types of windturbines and class five or higher regions are not commonly encountered.At fifty meters height, a class four average wind speed falls betweenabout 15.7 and 16.8 mph. Comparing this wind speed to that of typicalpropeller-driven aircrafts at optimal cruising speeds illustrates thegreat disparity between the most efficient speed for axial flow winddevices and the operational speeds encountered in tower-mounted windturbines.

Interestingly, in the operation of modern propeller-type wind turbines,both the cut-out speed (where blades are feathered to preventoverpowering the generator) and the park speed (where rotation isstopped and one blade is aligned with the tower in a self-preservationmode) have relatively low specific speeds. Much work has been done uponthe axial design using various airfoil profiles and blade rotation toachieve today's level of observed performance; however, the axial flowdesign is believed to utilize a much less than optimal flow orientation.

SUMMARY OF THE INVENTION

The present invention is directed to a multi-stage radial turbine forusage in energy capture from fluid streams with low to moderate specificspeed. The primary sources of low energy content streams include windand gravitational flow of water; however, other energy sources can beused. Additionally, the turbine can be driven and utilized as a liquidpump, gaseous phase fan, or gaseous phase compressor. The efficiency ofthe turbine allows it to be incorporated into many different devicessuch as, but not limited to, non-stationary land-based vehicles andwaterborne vessels. When the turbine is used for land-based vehicles orother types of non-stationary devices, the velocity differential betweenthe device and the surrounding fluid can create the motive forcenecessary to drive the turbine; however, this is not required. Theresulting shaft work produced by the turbine can be used to drivegenerators and/or other types of rotating equipment. The design andmechanism by which the turbine is driven is novel to the art. Thetraditional wind energy devices have air directed along the axis ofrotation of the turbine blade and such designs have been found to bemuch less efficient than the design of the present invention. It hasbeen found that applications for extracting energy from a fluid thatutilize wind or other low energy content streams stand to gainsignificantly increased performance by utilizing flow oriented in aradially outward direction. The radial design can also be optimized forhigher energy content streams where the initial stages' orientation isthat of transitional flow. These optimized configurations that benefitfrom initial transitional flow orientation typically will conclude witha radial section of stages that is far superior once the energy contentof the feed stream has been reduced.

In another non-limiting application of the present invention, combininga radial compressor, a burner and a radial energy capturing turbinecreates a new class of internal combustion engine that is analogous to amulti-staged, axial flow turbine engine. One non-limiting advantage ofthe radial design is that it can operate efficiently in flow conditionsthat are much slower and/or have rotational speeds that are greatlyreduced than the necessary minimum speeds for axial designs. Theseslower flows are actually most commonly encountered. For example,stationary fans typically are used to accelerate air that begins at nearstagnant velocity.

Of importance is the understanding of the physics-based mechanism of theradial turbine design that leads to huge performance gains overconventional designs. The radial turbine mechanism is “momentum vectordelta.” Conventional axial flow, propeller-style wind turbine devicesdirectly extract energy via impingement and/or lift generated by airfoilprofiles incorporated into the rotating blades. Both impingement andairfoil energy extraction can be considered as a high shear ratemechanism. The conventional axial flow designs are limited by the amountof energy contained in the fluid stream. The Betz's Law and its fluidstream energy content based derivation emphasize the energy contentlimitation of 59.3% of the kinetic energy of the wind stream.Conventional wind turbines achieve 75 to 80% of the Betz Limit.

The radial design is based upon capturing the force that is induced bythe stream's momentum vector being perpendicularly turned. This isanalogous to gravitational forces acting upon a satellite in circularorbit. Because the angle between force and fluid is perpendicular, thereis no work done on or by the fluid. The relationship between appliedforce and work is one of the cornerstones of basic physics and isexpressed as “the product of the component of the force in the directionof the displacement and the magnitude of the displacement.” Thefollowing equation describes this relationship:

W=(F cos θ)s

Where: W=work, F=Force, θ=angle, s=displacement

An applied force that is perpendicular makes the cos θ quantity zero.Thus, the energy content of the wind in the radial design does notdirectly contribute to the power and work being produced. The wind'skinetic energy is consumed strictly in overcoming the pressure dropthrough the system that is specifically optimized to keep pressure dropto a minimum. The effect of the radial turbine's vane curvature is tochange the stream's momentum vector direction and it is this change inmomentum vector direction (not magnitude) that creates the force uponthe radial turbine's vanes.

Regarding another analogy, consider an endless stream of footballplayers running by stationary lines of coaches arrayed in curved paths.Each coach pushes the player perpendicularly and changes the player'sdirection but does not affect the magnitude of the player's velocity.The radial turbine captures the energy expended by the arms of thecoaches. The amount of energy that can be extracted by momentum vectordelta can greatly exceed the kinetic energy content of the stream.

The new radial design turbine/collector combination of the presentinvention is believed to result in improved performance and whichexceeds the axial flow performance by an order of magnitude or greaterof prior devices. This huge performance difference can be observed usingthe provided graph (FIG. 2) and by simply visually extending the axialflow efficiency curve far to the left to where is located the very lowspecific speeds of fluid flow classification. As is evident from FIG. 2,the novel design of the present invention can obtain efficiency valuesthat were heretofore unachievable at low fluid speeds.

The predicted performance increase of the present invention results frommore than the handling of flow in a radial nature. Also of majorinfluence is the acceleration of the feed stream, the multi-stagehandling, and/or the precisely controlled discharge of the spent wind orexhaust stream. Mathematical modeling of the combined effects of themajor contributing influences and their associative relationships havebeen used to quantify performance expectations.

The typical arrangement for power extracting calls for inside-out flow.Power imparting applications mostly flow outside-in. In the case ofinternal combustion engines that utilize multi-layered, multi-stagedplatforms, the entrance and exit layers typically will adhere to theabove guidelines. Special circumstances that have spatial limitationsmay be encountered that require deviation from the typical flowdirection; however, performance degradation is expected.

Analogous to the stages found in axial flow turbines, radial turbinestages alternate between bending the flow to create a tangentialvelocity component and straightening the flow to align it in a radiallyoutward direction.

Most power extracting applications begin and end with a bending stage.The bending stages typically have a greater amount of curvature and thusare capable of extracting a greater amount of power than straighteningstages. Most applications will utilize stationary straightening stagesto simplify the mechanical complexity, reduce maintenance requirements,and/or reduce equipment cost; however, some applications with highenergy sources (such as compressed steam) may benefit from a counterrotating arrangement. Each application must be evaluated individually.

Power imparting applications are the reverse of extracting applications.The vane shape used to extract power by receiving a radially outwardflowing stream and bending it to induce a tangential velocity componentwhile allowing the stream speed to slow via expansion remains unchanged.However, its interaction with the stream is completely reversed in thepower imparting roll. Power is consumed as the stream is volumetricallycompressed and accelerated radially inward. Describing this action as“scooping” allows for ease of interpretation. The alternating stages'vane shape also remains unchanged. But again, their effect on the streamis also reversed. Further compression and acceleration takes place asthe stream is forced to take on a tangential velocity component. Thetangential velocity inducing stages are analogous to stator stages inaxial flow designs. Here, again, the stages with the least amount ofcurvature are typically stationary. The tangential angle of departure ofthe bent stream should be aligned with the entrance angle of thefollowing scoop stage. This alignment is illustrated in FIG. 14.

Mathematical modeling of the fluid dynamics has evolved enough to allowfor significant optimization and performance prediction of powerextracting applications. The model can be used to calculate and balancethe primary momentum transfer effects within the stages, the pressuredrop associated with each component of the system, and/or transitionaleffects from one component to the next. The model can be a bulk flowmodel that utilizes classical thermodynamics and traditional fluid flowequations (e.g., see equations in Crane Technical Paper No. 410).Optimization parameters include, but are not limited to: componentsizing; the number of stages; each stage's vane count, speed ofrotation, and curvature; the distributor geometry; the diffusergeometry; and/or the intra-stage height profile. Optimization generallyrequires a time-based performance determination over the full range ofoperational conditions. For wind applications, a Weibull wind speeddistribution can be used to fairly represent the time fractions atvarying weather conditions.

Each application can be optimized by primarily focusing on itsdominating characteristic. For low energy feed streams, the dominatingcharacteristic is generally pressure drop and thermodynamic balance isgenerally of secondary importance. The optimization of applications withhigh energy feed streams generally requires the design focus to beplaced on thermodynamic equilibrium and entropy minimization withpressure drop remaining important, but generally not being thedominating characteristic.

Further refinement of the mathematical model using discreet elementanalysis methods—advanced Computational Fluid Dynamics (“CFD”)—can beused to allow the inclusion of effects associated with secondary flowsso as to produce an improved performance prediction; however, this isnot required as secondary flow losses are included in the overallpressure drop equations found in bulk fluid flow type analysis. Theeffects of secondary flows are encountered any time a stream is forcedto change direction.

In one non-limiting arrangement of the invention, there is provided aturbine system comprising a collector, a distributor and one or moremulti-vane radial turbine stages. The one or more turbine stages areeither stationary or rotatable about a common axis. At least one of thestages is non-stationary and rotatable about the common axis. Thecollector is designed to capture and direct a portion of an open andfreely flowing fluid to the distributor. The distributor is designed toevenly distribute and turn the collated stream it receives into aradially outward flow that is oriented into the entrance of the a 1^(st)turbine stage. The fluid contact with a plurality of vanes as the fluidflows radially outward from the common axis causes said non-stationaryturbine stage to rotate about said common axis. The collector includes afirst opening, a second opening and a body passageway therebetween. Thefirst opening has a first cross-sectional area and is designed toreceive fluid and direct such received fluid into the body passageway.The second opening has a second cross-sectional area and is designed toallow fluid to exit the body passageway and the collector. The firstcross-sectional area is generally greater than the secondcross-sectional area; however, this is not required. A central flow pathof the collector body may remain in a straight line congruent with thedirection of the feed stream; or may have curvature of up to 90° toallow design freedom needed to place the distributor in the mostadvantageous orientation that is dependent upon both the specificapplication and size of equipment. The distributor includes a firstpassageway whose perimeter contains the fluid stream from the collector,a second passageway that directs the fluid stream into the first turbinestage, and an interconnected passageway body that contains the fluidstream between said passageways. The interior of the passageway body canoptionally include additional passageways that divide the primary flowinto multiple flows with parallel orientations. Thus at minimum are twosurfaces, one surface being an impact surface in regards that the fluiddirectly impinges upon its surface and a second surface being acontainment surface that completes the passageway body of saiddistributor. The effect of the distributor is to direct a collated bulkfluid stream into one that is radially dispersed and that directs theflowing fluid into the first turbine stage. The first turbine stage floworientation may be purely radial or it can be within the transitionalflow definition wherein flows are transitioning from being purely axialto that of becoming purely radial. A diverter valve can optionally beincluded that diverts away a controllable fractional amount of the bulkstream to limit the flow rate handled by the downstream radial stages. Aplurality or all of the stages can have an overall cylindrical geometry.One or more of the stages can include a first opening, a second opening,and a body passageway therebetween. Each of the openings can beprimarily oriented towards or away from the common cylindrical axis, andin the case of pure radial flow being oriented precisely towards or awayfrom the common cylindrical axis. The body passageway of the stages canbe confined by surfaces that comprise a floor and ceiling when thecommon axis is vertical with the horizon. The distance between the floorand ceiling can be the same at the entrance and exit; however, they willtypically have different measurements. Within the body passageway of oneor more stages is a plurality of curved vanes extending from the floorto ceiling and which interact with the fluid stream thereby causing saidstream's direction of flow to be altered. One or more of the stage'svane curvature is selected to perform one of two functions. Thesefunctions are to bend a radially aligned stream into having a tangentialvelocity component or straighten a stream with a tangential velocitycomponent into one that is radially aligned. The vertical vanes thatspan from the floor to the ceiling an optionally incorporate a cornerrounding profile, varying vane wall thickness, and/or varying floor toceiling height with one non-limiting purpose of maintaining a constanthydraulic radius along the flow path as used in low energy density feedstream applications or in the case of higher energy feed streams tominimize entropy delta or other thermodynamic objectives as necessitatedto optimize overall performance. The one or more stages that arenon-stationary can optionally have a suspension system that allowsindividual freedom of rotation around the common axis, and/or a commonsuspension system whereas one or more of the non-stationary stagesrotate as a combined unit, or a combination of combined and individualsuspension systems. The diffuser can include a first opening, a secondopening and a body passageway therebetween. The first opening has afirst cross-sectional area and is designed to receive fluid flowingradially outward from said outer most stage and to direct such receivedfluid into said body passageway. The second opening has a secondcross-sectional area and is designed to allow fluid to exit the bodypassageway and the diffuser. The first cross-sectional area canoptionally be less than the second cross-sectional area. Onenon-limiting purpose of the purpose of the diffuser is to smooth thetransition of the turbine system's exhaust in a manner that reducesoverall system pressure drop. A shroud can be optionally provided. Theshroud can be positioned near the diffuser to reduce interference fromfluid that did not enter the collector with fluid that is exiting thediffuser. The shroud may optionally have multiple sections around theperiphery of the diffusor to optimize usage of the negative gaugepressure draft that results from the fluid being displaced around anobject placed in its flow path. A pump, motor and/or electric generatorcan optionally be included. The rotation of the one or more turbinestages is designed to power the pump, the motor, and/or drive theelectric generator to cause electric power to be generated by theelectric generator. The fluid can be provided via a closed conduit inplace of the open collector. The one or more turbine stages can bedriven and consume power for the purpose of creating pumps, fans,blowers, and compressors. These driven devices will typically have flowdirection opposite that of power producing configurations. Thereforeflow is predominately radially inward. An increased stage count can becompactly achieved by having multiple layers of turbine stages. Flowwithin each layer of stages reverses direction from the preceding layerof stages such as from radially inward to radially outward or fromradially outward to radially inward. Additional body passageways canoptionally be added to direct one layer of stages outlet to the inlet ofthe next layer of stages. The turbine system can combine a turbinecompression of air, a combustion chamber, and a radial turbine powerextraction to create a radial turbine internal combustion engine. Thecompression and power extraction stages can be single or multi-layered.The turbine system can be mounted on a moving platform whereas thedifferential velocity between the platform and the surrounding fluidprovides the velocity gradient needed for the turbine system to extractpower. These platforms can be land based road vehicles (cars, trucks, orother wheeled transports), rail mounted vehicles (open rail trains orclosed rail courses used for power generation), or waterborne vessels(surface ships and underwater vessels). A movable frame can beoptionally provide that is designed to allow the collector, thedistributor, the one or more non-rotating turbine stages, and theoptional shroud to be connected to the frame. The frame can includewheels that enables the movable frame to travel on a road or railsystem.

The components comprising the new designs and their operationalcharacteristics, along with other features and advantages, will becomeapparent to those skilled in the art upon the reading and following ofthis description taken together with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

The drawings that are provided illustrate various embodiments that theinvention may take in physical form and in certain parts andarrangements of parts wherein:

FIG. 1 provides example radial turbine configurations for powerextraction (wind energy and hydroelectric), power imparting (pumps,fans, compressors), and combined power imparting and power extraction inthe form of a new type of internal combustion engine;

FIG. 2 is a graph that shows through curves of efficiency versusspecific speed whether the optimal flow regime is either radial, axialor the transitional range between these two;

FIG. 3 illustrates categories of non-limiting shapes of collectors inaccordance with the present invention;

FIG. 4 is a diagram that illustrates sized turbines where the maximumheight of each stage is equal as well as the minimum height of eachstage;

FIG. 5 is a diagram that illustrates turbine stage height relationshipswherein the maximum and minimum heights have been treated as variablesused to maximize performance without consideration to the verticalexpansion rate and the smoothness of flow from stage-to-stage;

FIGS. 6 and 7 are diagrams that illustrate the benefit of rounding thecorners in a collector;

FIG. 8 is a diagram illustrating a control volume analysis of the forcesgenerated by momentum vector analysis of a fluid flowing through aconduit with non-constant hydraulic diameter;

FIG. 9 is a diagram that illustrates the concept of a shroud that isused as a partial wrap-around shield to inhibit or prevent wind at neargrade level from interfering with the diffuser's upstream side radialdischarge. The shroud also allows the control and optimization ofnegative gauge pressure draft at the turbines/diffuser discharge;

FIG. 10 is a diagram that illustrates a non-limiting radial wind turbinedesign in accordance with the present invention as compared to aconventional, axial flow wind turbine system;

FIG. 11 is a graph that illustrates the benefit made possible by adownstream negative gauge pressure draft on an example wind energyconfiguration.

FIG. 12 illustrates a non-limiting and modified rail truck system inaccordance with the present invention. This truck system may be used tosupport and allow positioning into the wind collectors used in large,commercial scaled wind energy configurations;

FIG. 13 is a scaled drawing example of a wind energy turbineconfiguration with three stages: two rotating/bending stages and onestationary/straightening stage;

FIG. 14 illustrates the segments for both bending and straighteningstages as well as the impact and containment sides of the vanes;

FIG. 15 illustrates an example of secondary wall and corner roundingprofiles for a bending vane;

FIG. 16 illustrates a non-limiting residential or small-scale windenergy capture system with a 90° bend collector in accordance with thepresent invention;

FIG. 17 illustrates a non-limiting wind energy capture system with astraight walled collector shape for small-scale applications;

FIGS. 18-20 are graphs that illustrate power output for wheel-mountedwind energy capture systems. These performance graphs are based uponrectangular turbine cross sections;

FIGS. 21-23 are graphs illustrating performance in miles per gallon andenergy generation by wheel-mounted wind energy capture systems based onthe speed of the vehicle. These performance graphs are based uponrectangular turbine cross sections;

FIG. 24 illustrates a seven-stage single-pass blower wherein there arefour bending stages that rotate as one assembly and the threestraightening stages being stationary;

FIG. 25 illustrates the combining of a multi-layered inlet compressor, acombustion chamber, and a multi-layered outlet energy recovery turbineto create a radial turbine internal combustion engine;

FIG. 26 illustrates an arrangement for hydroelectric power generationwhere the dam height can be as little as a few feet in height;

FIG. 27 is a graph that illustrates the horse power verses vessel speedfor a submerged and forward facing straight shaped collector with aparticular waterborne vessel as an example case; and,

FIG. 28 provides two illustrations of different collector inlet shapeson a vehicle.

DESCRIPTION OF NON-LIMITING EMBODIMENTS OF THE INVENTION

Referring now in greater detail to the drawings, wherein these showingsare for the purpose of illustrating various embodiments of the inventiononly, and not for the purpose of limiting the invention, the presentinvention is directed to a multi-stage radial turbine for usage inenergy capture from fluid streams with low to moderate velocity. Onenon-limiting advantage of the radial design is that it can operateefficiently in flow conditions that are much slower and/or haverotational speeds that are greatly reduced than the necessary minimumspeeds for axial designs. Additional discussion and quantification ofexpected performance gains by controlling secondary flows are set forthin more detail below.

Power from Wind

In one non-limiting aspect of the present invention, the invention isdirected to a radial design having improved performance for producingelectrical power from wind-driven devices. Several of the design aspectsof the wind-driven turbine have a direct correlation to otherapplications. Thus, the wind energy discussion herein serves as a basisfrom which other applications can be derived with similarities anddifferences easily distinguished.

The components used to extract power from wind generally include: (1) acollector, (2) an inlet distributor that feeds the innermost turbine,(3) multiple radial turbines, (4) an exit or exhaust diffuser, (5) ashroud at near grade level to route prevailing wind around the turbineassembly, and (6) one or more generators with electrical conditioningequipment and controls.

1. The Collector

In the case of extracting energy from wind, the most visibly dominatefeature is the collector that captures a portion of the wind flow anddirects this flow to the inlet distributor. Four primary shapes for thecollector in accordance with the present invention are set forth in FIG.3; however, it can be appreciated that there are other shapes of thecollector that can be used in accordance with the present invention(e.g., usage of shapes with non-circular perimeter cross sections takenperpendicular to the primary flow path—see FIG. 27). One of thecollectors set forth is particularly favorable for large, commercialscaled power generating applications. It and the other non-limitingshapes illustrated in FIG. 3 are discussed as follows:

Shape #1: This collector utilizes an elongated entrance with a bendangle of about 90°. This is one non-limiting shape for equipment sizedto produce commercial quantities of electricity. This boxier shape, ascompared to Shape #2, provides greater inlet area and greater structuralstability characteristics. The sloped inlet floor allows for some of theprecipitation that may enter the collector to drain outward. Thelayout's radial lines can be equally spaced along a line connecting theendpoints of the cross section's lower curvature; however, this is notrequired. The resulting power output is potentially increased by greaterthan an order of magnitude based upon inlet area as compared against theswept area of current axial flow propeller designs. This is the mostfavorable shape for large scale wind energy configurations.

Shape #2: This collector incorporates a continuously reducing radius anda bend angle of about 90°. Essentially, Shape #2 is a wind sock shapewith a bend of about 90°. As can be appreciated, the angle of bend canbe less than 90°. The collector is capable of producing more than doublethe power of current radial designs when the collector's size is chosensuch that its opening's center is equal to the hub height of comparableaxial flow designs.

Shape #3: This collector incorporates a straight funnel or wind sockconfiguration. This shape has the advantage of no secondary flowconcerns. Small, residential-sized wind energy systems can be designedthat utilize this shape. This shape is also a good choice for waterborneapplications and applications where the inlet stream is contained withina conduit.

Shape #4: This collector utilizes a reducing radius and a bend angle ofless than about 90°. This shape is generally used for specializedapplications—typically vehicle mounted; however, this is not required.This shape allows for a transition from the straight funnel to any bendangle necessitated for practical reasons up to and including that ofhaving about 90° of flow directional change.

Collector shape #1 can generally be sized such that it has the mostcompact configuration for the amount of inlet cross sectional area whilemaintaining a circular cross section flow path perpendicular to averagevelocity vector from inlet to outlet. The larger opening is to beoriented directly into the wind stream and the bend is used to conveythe captured wind to near grade level; however, this is not required.Many advantages of near grade level can be achieved and these include,but are not limited to: (1) increased accessibility of rotatingequipment, (2) generators, electrical conditioning, and controls beingfixed at near grade level no longer have as strict of size limitationsas compared to tower-mounted axial flow designs, (3) no cable windingconcerns as a result of the generators being fixed in place, and/or (4)a reduced dependence upon high lift cranes needed for assembling andservicing. The collector can create a contained and accelerated flow.Internal baffling can optionally be used to minimize the twin helixsecondary flows that typically result from bending of primary flows.Available power varies approximately with the cube of fluid velocity andthe acceleration provided by the collector contributes to the overallperformance of the design. Approximately 60% of the wind directlyopposing the collector inlet may be rejected. Even with this largerejection rate, having a contained flow at increased velocity that isrouted through highly optimized radial turbine stages produces muchgreater power than axial flow, non-contained stream, propeller designs.

The radius of bend for the collector is generally at its maximum at thecollector's inlet where its diameter is the largest; however, this isnot required. The bend radius can continuously decrease as the collectordiameter decreases to its minimum at the collector outlet. The collectordesign can vary from that of a rigid structure (i.e., composed oftriangulated or geodesic beams as a framework) to one with greatflexibility (such as realized from choice of a single-walled wind sockor even a double-walled wind sock that is inflated to help maintain itsshape). One of the primary attributes of importance is the overallfunctionality of the collector in accordance with the present invention.

The collector may optionally incorporate an extended hood shape at itshighest point to minimize blowing precipitation from being entrained inthe captured airstream. Large, commercial-scale collectors can bemounted on tracks with horizontal alignment carefully achieved using astationary guide rail; however, this is not required. When used,baffling within the collector serves a dual purpose. Baffling allows thepressure drop associated with secondary flows to be minimized. Also, atwind speeds which push the generators beyond their capacity, the bafflescan be mechanically rotated to restrict the flow of wind through theentire assembly; however, one non-limiting method for adjusting flowrate can utilize a diverting valve configuration that is part of thedistributor and which is discussed in more detail below. The collector'sskin can be of rigid material such as a thin sheet metal; however, otheror additional materials can be used (e.g., plastic, composite materials,ceramics, sail cloth or specialized polymer sheeting may provide a moreeconomical choice of building material).

2. The Distributor

The distributor's purpose is to create the radial direction of flowwhile turning the fluid stream it receives from the collector such thatthe stream's flow is aligned with the entrance of the first turbinestage. For collectors with about a 90° bend, this radially outward flowcan be horizontal to the ground; however, this is not required. Thedistributor discharges into the first turbine stage that is typically arotating stage that bends the radial flow into a flow with a tangentialvelocity component; however, the first stage could be stationary andalso used to either radially straighten or condition the flow directionsuch that it optimally feeds the follow-up stage.

Current modeling of wind energy applications suggests that the inletdistributor is most likely a single cone-type configuration; however,other shapes can be used. The optimized diameter of the turbine stagescan be used to define the distributor's exit diameter. Modeling resultsemphasize the radial design's advantage as indicated by the downwardconcavity of the system's performance to distributor exit diametercurve. Even at wind speeds accelerated by the constricting process ofthe collector, the optimized configuration is solidly in the radial flowregime. This is emphasized in the cross sectional drawing of FIG. 4 thatillustrates the diameter relationships of the various post-collectorstream handling components. The adding of full or partial cones beyondthe single-cone configuration may improve the performance of wind energyapplications. Other power-extracting implementations using highervelocity streams are expected to trend towards reduced turbine stagediameters while still remaining radial flow in nature, but trendingtowards dimensions where peak power is achieved via higher rotationalvelocities. It is believed that the higher velocity feed streamapplications may have increased cone count, either full or partialcones, to help turn the stream about 90° in the more constrained spatialarrangement. Additionally, if the feed stream is of sufficient energy, ahighly optimized design might find the first and possibly even a portionof the second stage to have flow orientation within the transitionaldefinition. Transitional flow turbine configurations use the initialstages to perform some or even all of the feed stream directional changefrom a collated linear stream to being dispersed radially outward.Transitional stages utilize vane shapes that may be complicated by theamount of twist used to perform the additional directional change ascompared to purely radial stages.

The distributor can also provide the recommended location for thecontrol that limits the maximum quantity of wind allowed to the turbineassembly. This control can be a radial flow diverting-style valve thatboth splits away a portion of the wind stream and acts as an obstructionin the primary flow path; however, this is not required. The twodiagrams illustrated in FIG. 4 show the diverting valve in closed andopened positions. The plenum chamber exhaust may be ducted to providecooling air flow to the generators. The plenum is also a work space asit houses the diverting valve actuating equipment that will generallyrequire periodic inspection and maintenance. The positive pressure ofthe plenum suggests entryway doors to be of a sliding-type design forenhanced safety; however, this is not required. Additional analysis canbe used to quantify the amount of cooling air needed by the generatorsover the full range of operability. Slightly oversizing the collectorinlet to make up for the loss of diverted flow for cooling purpose canbe utilized. Additionally, the distributor can be incorporated into thecollector such that a continual decrease in cross sectional area occursup to the exit point of the distributor. This incorporation allows thethroat diameter of the collector to be slightly enlarged and allow for asmoother transition of flow between the collector and distributor.Interestingly, the usage of bulk fluid pressure drop equations resultsin collector/distributor wall angle and cross sectional arearelationships that closely duplicate those of an optimized venturi.

3. The Radial Turbines

The diagrams found in FIG. 4 illustrate turbines sized such that themaximum height of each stage is equal as well as the minimum height ofeach stage; however, this is not required. The stage height relationshipin this specific example uses a 1.5 to 1 ratio of maximum height tominimum height; however, other ratios can be used. Meeting therequirements of having a constant hydraulic radius and specified heightsinvolves a tuning task that is described more fully in the turbinesection set forth below. These stages can be tuned to have matchingbending stage RPM values throughout the entire operational range;however, this is not required.

The illustration in FIG. 5 emphasizes turbine stage height relationshipswherein the maximum and minimum heights have been treated as variablesused to maximize performance without consideration to the verticalexpansion rate and the smoothness of flow from stage-to-stage. The stepchange in height between stages is modeled as a flow restrictionorifice. Here a constant ratio of 1.1 to 1 of the stage outer radius toinner radius was maintained; however, other ratios can be used. Thus,the horizontal ring width of the stages varies as a function of stagediameter. Optimizing to performance only, without regard to physicalheights, can lead to some unwieldy and impractical configurations wherethe maximum height is more than double that of the previous example.

The FIG. 5 illustration shows that the bulk fluid model without physicalconstraints can converge to stage vane height delta values that can beunwieldy. This vane height delta is representative of the maximum amountof energy the fluid stream is capable of transferring to the rotatingturbine. The heights for this example were obtained by equating loss offluid momentum (e.g., stream velocity reduction) to the work performedwhile maximizing the later quantity. One non-limiting method ofspecifying the design is to constrain or force physical geometric limitsas part of the optimization task. For example, matching maximum heightand matching minimum height across all stages with a 1.5 to 1 ratio ofmaximum height to minimum height can be used as a starting point;however, this is not required. Optimization of this “squash” ratio mayrequire a combination of computational fluid dynamics and pilot scaleperformance measurements. Full sized production unit performance can beevaluated to confirm the accuracy of bulk fluid modeling, advancedmodeling, and piloting studies.

Secondly, the unmatched outlet to following stage inlet heights can beundesirable as it introduces an additional source of pressure loss. Eventhough the bulk fluid model includes the effects of the step change inheight between stages, the unconstrained optimization may bequalitatively inferior to a design that uses the smoothest possibletransition between stages.

Finally, as the wind speed changes, the RPM ratio between the twobending stages for both the constrained and the unconstrained designappears to follow a near perfect linear relationship. This linearity canallow constant ratio gearing to be used to couple stages that rotateindependently. The design can be further optimized to have the RPMvalues matched across the entire operational range and mount likefunctioning stages onto a common platform; however, this is notrequired. The usage of a common platform that reduces the design'smechanical complexity can be advantageous.

During simulation studies and optimization tuning, a particularlyimportant validation of the radial design was observed by keeping theouter to inner stage ratios constant and optimizing the first stageinner diameter for peak power output. Thus, the single-cone distributorouter diameter, the inner and outer radii of all turbine stages and thedimensions of the diffuser are spanned in a cohesive relationship whileplotting the power produced. The optimized result produces aconfiguration where the innermost turbine rotational speed is actuallyless than the outer bending stage's rotational speed. This reduced speedis a result of the first stage sacrificing its optimal diameter for theoverall combined performance of all stages. Even at the acceleratedfluid speed as produced by the collector's reduced outlet area, theoptimal diameter is in the purely radial regime. It is not transitionaland certainly far removed from axial flow.

Generally, the most efficient radial turbine arrangement for powergenerating applications is having the feed to the first stage flowingradially outward; thus, the first stage is the one that bends the streamsuch that the stream has a tangential velocity component upon exitingthe first stage; however, this is not required. It is possible tocondition the stream prior to the first stage by adding a stator toimpart a tangential velocity component and thus the corresponding firstrotating stage would be a straightening stage; however, only a verypeculiar type of application would require such an arrangement and suchconfiguration may not be desirable. Therefore, the first turbine stageis generally not constrained to only being a bending stage, butmechanical efficiency concerns strongly push the design towards havingthe first stage as a bending stage. The functional purpose of the stagesshould alternate between bending and straightening; however, this is notrequired. For example, the placement of two stages adjacent to eachother that both bend the stream away from a radial flow direction wouldin affect perform the same task as a single stage which used vanes withgreater curvature. Bending stages generally have greater curvature, aregenerally more efficient, and generally capture a greater amount of thewind energy than straightening stages. There are several dimensionalparameters that can be optimized for each stage. These parametersinclude, but are not limited to, the stage depth or difference betweeninner and outer radii, the vane count, the curvature of the vanes, andthe height profile. The current model simulates the primary wind tomechanical energy mechanism using segmental analysis; however, othertypes of modeling can be used. Each vane's concave side acts as thedirect impact side that bends or straightens the fluid stream. Thevane's convex side is generally the containment side (see FIG. 13). Thecurrent simulation and computer modeling utilizes momentumrelationships, bulk fluid flow pressure drop calculations, andrectangular cross section turbine passageways; however, this is notrequired. The turbine passageways have been simulated by dividing upeach passage into segments and calculating center of mass, angularmeasurements, flow, pressure drop, and work for each segment. Thesegments used in mathematical simulation are created using lines thatdivide the vane area into equally spaced lateral spans. The resultingsegment is a truncated pie slice shape bound by the opposing convex andconcave surfaces of adjacent vanes. The turbine's purpose is theconversion of force from the stream's change of momentum vector intorotational work. Recognized negative effects of the turbine include theparasitic friction between the rotating stage and the airspacesurrounding the turbine and the negative work between the bending stageexiting segment and the same stage's outer perimeter. This negative workis a result of the stream being pushed by the containment vane beyondthe zone where the impact vane is extracting power. The negative workamounts to approximately 2% of the positive work that is captured and isincluded as part of the bulk flow simulation.

The turbine flow pathway can potentially be improved by incorporatingoptimized vane height profiles and/or the use of baffles to minimizesecondary flow effects within the turbine. An expected additionalbenefit of profile enhancement and usage of baffles is that the vanescan become structurally stiffer. Increased stiffness can help reduce thelevel of mechanical strain on the rotating assembly and allow higheroperational rotating speed. Entry and exit vane profiles of both stagetypes can be a narrow, vertical, and/or slightly rounded knife-edgedshape to minimize the pressure drop associated with stage-to-stage flowtransfer; however, this is not required. Non-rectangular turbine crosssection profiles can be achieved by a gradual rounding of therectangular corners beginning at the stage entrance combined with agradual squaring-up of the rounding such that rectangular cross sectionis achieved at the stage exit.

Further enhanced control over the expansion outside the segmented zonecan be achieved by adjustment of the containment wall thickness. Thelocation of increased thickness of the containment vane for control ofexpansion can also provide a possibly desirable location for addingbetween vane wall structural stiffening.

Each application's optimization generally requires careful thermodynamicmodeling. The choice of the best thermodynamic model is generallydependent upon the energy content of the feed stream. For streams withrelatively low energy content (such as wind), the pressure drop is thedominating parameter and the model of choice is generally one thatmaintains constant hydraulic radius along the flow path within eachindividual stage. The varying vane count, radii, and other geometricquantities between the stages almost always demands that the fullyoptimized design also have different stage-to-stage values of hydraulicdiameter Applications utilizing higher energy content streams maybenefit by profiles that aim to minimize entropy change or otherthermodynamic objectives. All of the models may benefit from the usageof computational fluid dynamics.

Although the rectangular cross section bulk fluid model allows for aconsiderable amount of optimization, a more comprehensive understandingof the hydraulic radius effects may be gleaned from its definition, somesimple examples set forth below, and the below discussion about theeffects on turbine stage power output.

Hydraulic Radius formula:

$R_{H} = \frac{{cross}\mspace{14mu} {sectional}\mspace{14mu} {flow}\mspace{14mu} {area}\mspace{14mu} \left( {{sq}.\mspace{14mu} {feet}} \right)}{{wetted}\mspace{14mu} {perimeter}\mspace{14mu} ({feet})}$D = 4R_(H  {Equivalent  Diameter})

With reference to FIG. 6, a square with one unit side length has thesame hydraulic radius as a circle with unit diameter. However, the crosssectional area of the circle is obviously the lesser and thus thevelocity through the circle for a given mass rate is faster. Theresidence time of a segment affects the power calculation through itsmathematical influence by being in the denominator. Faster velocity andreduced cross section both decrease the residence time. A smallerdenominator equates to a greater amount of power generation by having alower Δt (residence time).

With reference to FIG. 7, further understanding of the benefit ofrounding the corners is achieved by considering a rectangle with aheight twice that of its width to an equally sized shape that hasrounded corners.

The rounded corner shape has a larger equivalent diameter and thus areduced pressure drop and greater mass flow rate is realized. Therounded corner's reduced cross section generally produces a greatervelocity and reduced residence time. Again, rounded corners can lead togreater power production.

Calculation of the power produced is generally a three-step process:

1) Calculate the reactionary forces imparted on the vane from the motivefluid. An example of reactionary force analysis and resulting equationsis found in FIG. 8 and is taken from “Fundamentals of Momentum, Heat,and Mass Transfer”; authors: Welty, Wicks, Wilson; ISBN 0-471-87497-3.The relationships between conserved forms of energy, non-conserved formsof energy, and Newton's second law are applicable in the analysis of theradial turbine.

2) Calculate the amount of work performed:

work=force*distance

The work calculation requires adjusting the two planer reactionaryforces from step #1 such that their component contributions are alignedwith the center of mass of the segment being analyzed and to that of aradial line that originates at the center of the circular turbineassembly. Also required is calculation of the distance the center ofmass moves during the defined time span. The amount of work performed isalso positively influenced by the rounding of corners that decreases thecross sectional area and volume of a segment.

The time span in seconds:

Δt=segment volume(cubic meters)*average density(kg per cubic meter)/massrate(kg per second)

Distance is a function of the turbine's rotational speed:

Sweep angle(radians)=RPM/60*2*Pi*Δt

Radius: r=distance (meters) from rotational center to segment's centerof mass

Distance=Sweep angle*r

3) Calculate the power produced which is the time rate of work done.Here, as well as in the previous step, residence time enters into thecalculation. The previous mention of the advantages of using roundedcorners is quantified in this power calculation step.

Power(Watts)=work/Δt

Vane Curvature

The layout of the curvature of the turbine vanes is accomplished byusing a spiral equation. One such spiral equation makes use ofconvenient polar coordinates:

r=ae ^(bθ)

The careful specification of constants “a” and “b,” the starting angle,the amount of angular curvature, and both x and y offsets from theorigin leads to a perfectly behaved curve where both ends meet targetedangular values to that of a radial line from the common center and whilesimultaneously matching the inner and outer radii of the turbine stage.FIG. 13 illustrates proper vane curvature where the entrance of bendingvanes and the exit of straightening vanes are aligned with a radial lineoriginating in the common center of all stages. Additionally, FIG. 14illustrates that a vane's exit angle for any particular stage shouldmatch the next encountered stage's vane's entry angle. Circular arcs orelliptical arcs may not provide sufficient control; however, by settingthe “b” constant to zero, a circular arc is achieved. If the “b”constant is greater than zero, an increasing radius spiral is created.By choosing a negative “b” constant value, a decreasing radius spiral isachieved. The polar equation has been found to work properly. There maybe steric limitations associated with the spiral curve. The vane countcan also be an important part of the layout and an insufficient numberof vanes can create geometric conflicts.

A Weibull distribution wind speed and the 1/7 th power law have beenused in the bulk flow computer simulation used for predictingperformance of the proposed design. Accurate performance predictions aredesirable to fairly evaluate the mobile applications that are discussedin more detail below. The 1/7 th power law allows integration of thewind speed over the height of the collector's inlet. The power lawprovides the adjustment to wind speed for heights other than themeasurement height which is typically about 30 to 50 meters; however,other heights can be used. Current mathematical modeling places thedesired number of stages at about three for the energy from windapplication; however, this is not required. The increased pressure dropcaused by adding stages beyond the optimal produces a greater collectorrejection rate that results in an actual decrease in overall power thatcan be achieved.

The turbine bending stages are expected to be mounted upon a supportingframework that is suspended and allowed to rotate; however, this is notrequired. The supporting framework may be incorporated into the turbinestage design itself and is not necessarily a separate entity. There areat least three possible options to suspend the platform; however, eachmanufacturer will need to evaluate all options and choose a method ofsuspension that best suits their specific design objectives. Aspreviously mentioned, the stages can be independent and rotate atdifferent speeds or they may have their geometry carefully matched suchthat their rotational speeds are also matched which allows a singlerotating platform to be used for multiple stages. The reduced mechanicalcomplexity and added rigidity resulting from usage of a single platformmakes the usage of matched RPM stages potentially desirable.

Suspension Method #1: The platform incorporates on its underside acircular rail with either rounded or tapered contours that are used as acontact surface with rollers that are mounted in a fixed location. Therollers can be grooved or shaped to receive and cradle the circularrail. This is one of the simpler design choices. The load on thebearings in the roller hubs is generally greater than that of method #2.

Suspension Method #2: The platform incorporates an underside circularrail similar to that of method #1; however, here the rollers ride on asecond circular rail that is positioned and fixed in place under therotating framework. The fixed rail is of different diameter than that ofthe platform such that there is an angle to the vertical through therollers' centrally located bearings. This angle or tilt of the rollersprovides for a self-centering action where the weight of the turbineassembly maintains its lateral position. The roller bearings can beconnected to rigid spacers that maintain equal roller spacing. One pairof rollers can be left without an interconnecting spacer to allow forslight expansion with changing temperature; however, this is notrequired. The load on the bearings can be greatly reduced as compared toMethod #1. The bearing load can be reduced to only that needed tomaintain spacing.

Suspension Method #3: Usage of linear motors. A roller assembly such asMethod #1 or Method #2 can be used. At speed, the linear motor lifts therotor assembly such that roller contact is broken. This magneticlevitation is analogous to high speed train operation. The linear motoralso serves as the generator and eliminates the need for gearedtransmission of force from the turbine assembly to the stationarygenerator(s). This option should be the quietest; however, even theroller suspended methods are not expected to create unreasonable amountsof noise. Current state-of-art linear motors generally are not asefficient as highly optimized conventional generators housed incylindrical framework. This suspension method is worthy of considerationas it would reduce long-term maintenance tasks on rollers and rollerbearings; plus the linear motor generator design has no mechanical ballor roller bearings as well.

The mechanical transfer of power from turbine stages to generator(s) canbe accomplished in several ways. One non-limiting method is through theuse of a large segmented ring gear mounted to the undercarriage of eachrotating assembly. If independently rotating stages are being used,transfer or idler gears can be used to transmit power from one stage toanother. In one non-limiting design, the use of three or more equallyspaced generators allows the radial gear force vectors to cancel out andmaintain the rotating assembly in a centered position (see FIG. 12). Thechoice of gear tooth pattern and the reactionary forces needs carefulconsideration. The straight cut gear pattern can be used for efficiencyand its pure radial reactionary force that avoids the non-straight cut(e.g. helical cut gear pattern) gearing axial reactionary force that canlift the turbine. No gear slap noise as a result of gear backlash isanticipated due to the generator(s) providing a continuous, althoughfluctuating, source of resistance. Optional hold-down rollers can beused. The hold-down rollers, when used, can be frame mounted and bepositioned immediately above a suitable surface of the rotatingassembly; however, this is not required. The hold-down rollers can beused as insurance against the rotating assembly lifting in the event ofan earthquake or unforeseen reactionary forces that may only becomeapparent at the higher operational speeds. The proposed design isexpected to have exceptionally low sound level of operation due to thesmooth and gradual slowing and expansion of the captured wind stream.However, it should be noted that attention to gear and bearing noise andits transfer through the turbines should be considered. The mostlyincreasing flow path cross section after the diffuser inlet might have amegaphone-like behavior. Reasonable usage of sound-dampening elastomericrubber isolation between turbine and ring gear should produce very lowsound emissions. Additional sound reduction can be achieved using soundabsorbing curtains in the wheel house (the area where the turbines aremounted) as well as sound dampening material applied to the outersurface of the floor of the turbines. The expectation is that soundlevels of operation may be well received by the public and no lowfrequency noise that plagues the axial flow propeller designs isexpected. Routine sound measurement and analysis should be part ofnormal operational procedures as any change in the sound profile mightbe indicative of a need for mechanical repair.

4. The Exhaust Diffuser

A component that can significantly impact the overall efficiency is theexhaust diffuser. The bulk flow simulation model used in the developmentof one non-limiting design of the present invention uses a pair ofparallel disks to approximate the performance of an optimized bellshaped radial diffuser. The diffuser pressure drop is adjusted byincreasing and decreasing its outer diameter. These changes in diameteraffect the exiting cross sectional area and the contained stream's exitvelocity. A more advanced bell shape is expected to further enhanceperformance; however, other shapes can be used.

5. The Shroud

The shroud can be a partial wrap-around shield that inhibits or preventswind at near grade level from interfering with the diffuser's upstreamside radial discharge. The shroud also can include multiple sectionsthat each help scavenge away the discharge stream. One such shroud isillustrated in FIG. 9. The shroud's shape and size can be optimizedusing CFD; however, this is not required. A simplified visualization ofthe shroud is to think along the lines of a shape reflecting the fronthalf of a racing bicyclist's helmet or a pair of streamlined sunglassesworn by athletes. The shroud can be attached to the assembly thatsupports and allows rotation of the collector. Thus the shroud alignmentcan be maintained in relationship to the collector's inlet.

The shroud can also enhance and control the downstream pressurereduction or draft that is created by wind flowing around an object.Negative gauge pressure draft can have a dramatic effect on the quantityof power produced. The relationship of power to draft is generallylinearly behaved. FIG. 11 provides a graph that illustrates the effectsof draft pressure on the amount of electrical power produced.

6. The Generator(s)

The manufacturer generally specifies the number, orientation, location,and capacity of generators utilized to convert the captured mechanicalenergy into electricity. In one non-limiting configuration is the usageof three vertical axis generators equally spaced around the perimeter ofthe rotating assembly with the highest tip speed; however, otherconfigurations and/or numbers of generators can be used. The radialdesigned wind energy layout provides great flexibility of generatorcombinations. The degree of this flexibility generally is not realizedin axial designed wind turbines. Cable winding is also not an issue withthe radial design because the generators can be fixed in place and neednot rotate with changing wind direction. Maintenance, servicing, andinstallation tasks are all made significantly easier as compared totower mounted axial flow equipment designs that place the generator(s)in confined nacelles located high above grade. These same advantages arerealized in conjunction with Suspension Method #3 as previouslydiscussed.

Performance Prediction and Computer Simulation

The description of the invention to this point is aimed at introducingthe principal components of the radial wind turbine design where theentire structure is stationary. Some non-limiting new and excitingapplications can involve having the radial turbine mounted on land basedvehicles or waterborne vessels. To enable performance predictions of themobile configurations, the performance of a stationary design isgenerally established as a reference point. The following simulatedperformance is believed to represent a conservative, below realizedexpectation. The conservative estimate is used to discuss mobileapplications so as not to exaggerate their expected performance. Anyperformance values which are inclusive of an estimated negative orpositive contributor will have the contributor(s) specifically itemized.One non-limiting configuration of the design is illustrated in FIG. 10and compared against a conventional, propeller style wind turbine. Thepropeller's diameter and the diameter of the collector inlet areillustrated as being similar for this non-limiting example.

Both equipment cost estimates and performance data are desirable toproperly size the generators and determine the operational capacity. Themobile applications, yet to be discussed, generally requirequantification of the expected performance to establish a basis orreference point. Generator sizing is best based upon an economicanalysis and typically not one of meeting a specific capacity factor,thus the usage of economic data is desirable to fairly size thegenerator(s) of the radial design and to determine the averaged expectedoutput or capacity. Cost data and sources are as follows:

http://www.windustry.org/resources/how-much-do-wind-turbines-cost 1.75Million $ per MW installedhttp://www.repp.org/articles/static/1/binaries/WindLocator.pdf GearboxGenerator & & Power Nacelle & Drive Elec- Rotor Controls Train tronicsTower Total Com- 28.0% 21.7% 17.3% 7.0% 26.0% 100.0% ponent Cost % $/MWComponent Installed Generator & Power 122,500 All Other Hardware1,627,500

After a reasonable cost model is established, it is the shape of the ROI(return on investment) curve that is of primary importance foraccurately sizing the generator capacity. The peak of the curve locatesthe optimal generator size while the magnitude of the ROI can be ignoredwhile the focus is on feasibility of design and not that of making abusiness decision.

The non-limiting economic model used for analysis herein produces agenerator capacity of about 40 MW and a capacity factor of about 0.3038.Axial designs have a typical capacity factor of about 0.34. Thecomparison unit has about a 1.6 MW rating. The axial flow comparisondesign thus produces an averaged expectation of about 0.544 MW. Thisvalue contrasts dramatically with the radial design averaged expectationof about 7.934 MW. The power from the radial design is estimated by thecurrent bulk flow model to produce about 14.58 times that of an axialflow design. Also, the axial design used in this comparison has aslightly larger swept area as compared to the base case radial design'scollector inlet.

The simulation results have been downgraded by 15% as a conservativelylarge estimate of the combined inefficiencies of the generators,mechanical gearing, and parasitic drag associated with the rotatingturbines.

Advanced CFD is desirable to accurately predict the enhanced performanceof internal baffling in the collector to minimize pressure drop effectscaused by secondary flows, to predict the draft pressure an optimizedshroud design can generate, and to fully explore non-symmetrical andnon-rectangular cross sections in the turbine passageways. The followinglist identifies missing contributors that shall have a positiveinfluence on expected performance.

Positive Contributors Missing from the Model:

1) An optimized shroud design that produces a negative draft pressure isexpected to dramatically enhance performance.2) The collector is the largest source of pressure drop through theentire system and significant improvement is expected by optimizingbaffles that minimize the effects of secondary flow.3) The usage of rounding corners and advanced turbine path optimizationis expected to provide enhanced performance.4) The bulk fluid flow equations are typically considered asconservative and overestimating of pressure drop.

Although the current bulk flow model does not accurately predict theamount of draft the shroud will produce, it does allow the draft to bespecified and generate expected performance at that assumed condition. Aplot of this output as illustrated in FIG. 11 shows that the presence ofdraft influences performance greatly on a system dominated by pressuredrop. Additionally, any reductions of collector pressure drop throughusage of baffling is expected to behave generally linearly and be ofabout equal magnitude effect as that of the draft effects. In otherwords, a 5 Pascal reduction in pressure drop through the collector isequivalent to about a 5 Pascal draft at the system outlet.

For example, at a wind speed of about 10.4 meters/second, the pressureat the collector inlet is approximately 105 Pascal. The collectorpressure drop is approximately 72 Pascal. A 20% improved collectorpressure drop combined with a draft of 20% of the inlet pressureproduces an approximately 35 Pascal pressure differential gain. It isreasonable to thus expect a performance improvement. Possibly a 40% ormore improvement can be realized. However, estimated performance graphsand values used in this patent application reflect the base case withoutthe effects of fully optimized collector, vane profiles, and shroud.

Collector Positioning

The large size of the collector that rotates into the wind and maintainsalignment with the distributor, turbines, and diffuser can provideunique challenges. To accomplish this task and maintain precisealignment, an installation of a rail system may be used. Onenon-limiting rail system is illustrated in FIG. 12. Train or crane railshave a range of profiles and linear weights that correlate to their loadcarrying capacity. The overall weight of the collector and itssupporting space frame will generally determine the best choice of rail.Some modification upon conventional twin-axel four-wheeled rail trucksmay provide sufficient carrying capacity and positioning capabilities.Additional rollers may be needed to align the truck horizontally uponthe concentric circular rails. These rollers can be used to replace thewheel flanges as the control over the tapered wheel's profile point ofcontact with the rail. Matching inner and outer wheel contact points andwheel dimensions can allow for efficient solid axel utilization where asingle axel revolution can cause each wheel to cover the same percentageof its respective total rail length. The outer wheel path covers thegreater distance. Slight shimming of the axels may be used to align eachto a common center of the circular track and reduce wheel-to-trackcontact slip to a minimum for this specialized service where the onlydirection of travel is circular. The truck design can be furthermodified through elongation. This design can achieve greaterdistribution of load bearing contact with the space frame and allows thetip-over clips (as discussed below) to spread their occasional loadsover a longer length of rail.

The structural design wind speed can include tipping effects of thecollector assembly. Even though the design provides for a wide base, theoverall weight and center of gravity should be evaluated to assuretip-over does not occur, even in the event that control that maintainsthe collector facing directly into the wind is lost. Tipping at any windangle for the design speed is generally to be avoided or controlled. Theadding of ballast can be considered to increase the wind speed attip-over rating; however, other options can be used. For example,stoutly mounted and adjustable retaining clips that only make contactwith the rail upon tipping initiation can be used. These clips shown inFIG. 11 can be used to maintain a minimal clearance from the railsduring normal operation. The clips could utilize rollers to make railcontact; however, the collector assembly is not expected to be rotatedduring a rare storm-induced tipping event. Thus, the clips can beprofiled to match the underside of the rail head to maximize contactarea. The rails can be secured sufficiently to counter the upward stressencountered during a tip-over event. A non-ballast rail system can beused and a base that allows the rails to be connected (e.g.,skip-welded) in place. The number of clips, their spacing, and theamount of welding needed to accomplish the design objectives is variableand specified according to the design's wind speed. Conventional railretaining clips may suffice without the need for welding, and possiblyno tip-over constraint may be used to meet the design objectives.

Secondary Flow and Additional Turbine Discussion

Turbulence and secondary flows in streams of fluids represent lostenergy and, when possible, should be minimized. In addition to theradial design taking advantage of the natural volumetric expansion withincreasing diameter; there are also localized and undesirable volumetricexpansions at the exit of bending stages and the entrance ofstraightening stages when the stages are defined by a simple constantwall thickness vane. A fully optimized design may require vane shapesthat are more complicated than having a constant thickness wall withprecisely specified curvature and height profile. Economicconsiderations may supersede the usage of complex vane shapes.

A square shaped stage entrance and exit is one non-limiting shape;however, a square shape is generally not encountered along the flowpath. Optimization can include finding the best vane count per stage. Asvane count is increased, a desirable increase in the effective bendangle is generally realized. Countering the improved bend angle is anincreased pressure drop from increasingly greater deviation from theideal square shape. As the rectangular width-to-height ratio deviatesfurther from the square 1:1 ratio, the decreased hydraulic equivalentdiameter results in increased pressure drop. Any pressure drop increaseresults in less mass flow rate from which to extract energy.

The dimensions and vane counts of a non-limiting and matched set ofstages shown in FIG. 13 are representative of the base case collectordimensions. The radial design scales nicely; however, the precisegeometry should be optimized for each specific application.

FIG. 14 illustrates the mathematical layout of segments used in theanalysis of the radial turbine design. Calculation of the wind force onthe vanes is analogous to calculating the force exerted upon a reducingpipe bend. To facilitate the calculation, ten equally spaced divisionswere made along the impact side of the vane; however, this is notrequired. The vane's shape is described using a modified spiral equationthat allows precise specification of the entrance slope, exit slope,entrance location, and exit location. The spiral equation is wellbehaved. It produces a smooth and continuous curve similar to a circle;however, the spiral has a continuously varying radius. A very narrowslice trapezoidal numerical integration method provides accurate volumeand center of mass determination of each segment. Energy is beingtransferred from the stream all along the flow path within thedivisions. The height profile of each stage is a direct consequence ofthe dominating characteristic of the system. For streams of low energy,the dominating characteristic is generally pressure drop and this isgenerally minimized by specifying constant hydraulic radius along eachturbine flow path.

The straightening stage will typically be stationary and the segmentalanalysis is again used to assure the hydraulic radius is equal along theflow path.

Two other recognized shape-tuning options to alleviate undesirablelocalized expansion and maintain stream velocity are corner rounding andusing variable vane wall thickness or a secondary wall as illustrated inFIG. 15. Corner rounding has the benefit of transitioning towards a moreefficient circular cross sectional shape. Thus the wetted perimeter isreduced, pressure drop is slightly reduced, and mass flow rate isslightly increased; while maintaining the stream's velocity andminimizing undesirable expansion between the last containment segmentand the stage's outer periphery.

Within the segmented region of the turbine passageways, the corner fillprofile is interpreted as a ratio of the distance from the vane surfaceto the opposing impact or containment vane. For example, when theprofile and the centerline coincide, the quarter circle fill has aradius of half the distance between the vanes as measured congruent withthe lines defining segment boundaries. Outside the region where a linecotangent to the centerline intersects the impact vane, the beforementioned interpretation loses validity. In the outer region, thecombined effects of the corner fill profile and the secondary wallthickness become specified by distance from the convex containment vane.Although a secondary wall can be incorporated with the concave impactside of the vanes, this may not be desirable as it would alter theperfectly behaved spiral geometry of the impact surface.

The flow straightening turbine stages utilizing strictly rectangularcross sectional areas also can exhibit undesirable expansion. Thestraightening stages expansion occurs at the inlet as opposed to theoutlet of bending stages. Again, utilization of a secondary wall andcorner filling help minimize any undesired expansion. Controlling theexpansion can effectively increase the equivalent hydraulic radius,reducing pressure drop, and/or maximizing mass throughput and/orinternal stream velocity.

Residential or Small-Scale Wind Energy Capture

The radial design is amenable for small-scale applications as well.These may include private residences or businesses located in favorablywindy regions. Both collector Shape #1 with the generator mounted neargrade as illustrated in FIG. 16 and the straight wall collector, Shape#3, as illustrated in FIG. 17 with an inline generator are feasible.

Collector Shape #1

A non-limiting wind energy system illustrated in FIG. 15 includes: theuse of Shape #1 for the collector, a shroud, and unifying framework thatprovides the structure to unify or tie a majority of the mostlystationary components together. These components include: both upper andlower halves of the distributor, the diffuser, the straightening stage,the collector, and the tail assembly. Note that baffles can be used inthis design within the single cone distributor to allow the lower halfof the distributor to be fixed in place through the upper half that isdirectly frame mounted.

The necessary steering tail assembly configuration can be quite varied.It may consist of a twin tail configuration, a circular tail, a largesingle tail, or other shapes that serve the purpose of aligning thecollector inlet into the wind. To maintain the collector facing into thewind's predominate or averaged direction of flow, a friction typebearing or viscous fluid dampener can be used to control any oscillatorybehavior of the turbine assembly caused by the ever presentinstantaneous wind directional changes. Counterbalance weight should beconsidered to offset the tail assembly weight and maintain the center ofgravity in line with the tower support bearing.

The small scale units can have control over the maximum internal windvelocity. This can be achieved by a single method or a combination ofseveral methods. Among the choices for control are:

1) steer the collector inlet away from being directly in line with thewind through the usage of yaw type controls that use tail assemblyposition to achieve the offset angle required to keep internal windvelocity within specification;

2) use secondary flow control baffles within the collector that aredesigned to pivot and obstruct a portion of the flow;

3) design the vanes of the shroud to pivot and obstruct the flow on thedischarge side of the turbines to create sufficient pressure drop tolimit flow rates as needed;

4) use the brake to park the turbines during excessive wind events;and/or

5) oversize the generator such that generator load can keep turbine RPMwithin tolerance.

Each specific design should have the drive shaft properly sized. Inaddition to end bearings, bearings along the drive shaft length areexpected. Particular drive shaft detail may also be needed for eachdesign such that assembly and maintenance tasks are not made undulydifficult. Spline type couplings and tower access ports may be used.

The tail assembly and collector generally have a greater amount ofallowable flexural tolerance than the stationary and rotating assembliesthat combine to create the wind pathway downstream of the collector.There generally should not be excessive gaps between the turbine stagesand during all wind conditions the individual turbine stages must notmake contact with each other. Precise component alignment during bothmating and separation tasks between the rotating turbine stages and thestationary stage(s) generally is maintained during the entire process toavoid damaging the turbines. Guide rods can be attached to the frameworkand allow tight tolerance bushings located in the turbine base plate tohelp maintain the alignment. Additional means beyond guide rods andbushings for keeping the turbine stages radially aligned and axiallyparallel during the assembly and disassembly tasks can be provided.

Straight Collector

Another non-limiting wind energy system illustrated in FIG. 16 utilizescollector Shape #3 along with a list of components the same as theprevious example. The straight collector shape for small-scaleapplications can provide better overall flow characteristics and producemore power. However, the straight configuration is generally onlypractical for relatively low power capacity designs. As the size of thesystem is increased, horizontal flow turbines may become impractical.Large horizontal axis radial turbines can have difficulty providingadequate structural integrity for the turbines, can have rapidlyincreasing tower requirements, and/or can provide more difficultequipment access. The cable winding issues for small-scale units can beremedied by using a slip ring connection.

As with every radial turbine design, the tasks of assembly anddisassembly require proper alignment to be maintained during the timethe rotating and stationary assemblies are within close proximity toeach other. The platform can be designed to provide a machined channelto facilitate these tasks. The channel, when used, can be designed toreceive the bearing housing of the rotating turbine and allow it to beevenly slid into place or evenly extracted. Design considerations areneeded to prevent rust, scale, and debris from insects from fouling thechannel. A sufficiently long channel can allow a precisely machinedpush-block to contact the bearing housing and prevent turbine stagecontact during these tasks.

The generator base can also use this channel and thus simplifygenerator-to-turbine hub alignment. Possibly, the generator base canserve as the above mentioned push-block; however, this is not required.

Any component alignment issues also extend to the framework that tiesall the components together. The framework is a component that shouldwithstand the full range of wind conditions without excessive flexing.The framework also should be robust enough to withstand handling duringshipping and lifting onto and off the tower. Specifically, engineeredlifting lugs and support cradles can be incorporated into theframework's specifications.

Mobile Applications

The greatly improved performance that the radial design provides forstationary wind energy applications can also be extended to mobileapplications. Both power produced and power consumed by aerodynamic dragcan increase with increasing speed. Of great interest is the observationthat the power produced can outpace the drag induced losses. If avehicle is fitted with a collector of sufficient size, there is avelocity at which the wind energy potentially matches the combinedrolling resistance, the drag of the vehicle, and the drag of the winddevice itself. Above this self-sustaining speed, excess power canpotentially be produced which can be used to recharge batteries or becontrolled through a throttling mechanism. If the vehicle is railmounted, the excess power can be transferred to the electrical supplygrid using either a 3rd rail mechanism or linear motors. A circulartrack arrangement would allow multiple, equally spaced vehicles tocontinually produce power. Capacity can be increased by track lengthwith more vehicles and by adding multiple concentric circular tracks.The vehicles on odd numbered tracks can be designed to travel inopposite rotation of those on even numbered tracks to minimize followingturbulence. The realized inlet speed might be less than the vehiclespeed if air is being dragged along by the leading vehicle. More tightlyspaced vehicles can be expected to have to travel faster to create agiven amount of power compared to widely spaced vehicles.

Values generally needed to calculate the combined aerodynamic drag androlling resistance include air density, relative velocity of vehicle toair, drag coefficient, rolling resistance coefficient, cross sectionalarea, and the normal force of the vehicle to the surface.

F _(D)=½ρv ² C _(D) A

http://en.wikipedia.org/wiki/Drag_equation

F=C _(rr) N

http://en.wikipedia.org/wiki/Rolling_resistance

The rolling resistance should be included in the estimation; however, itis a small contributor and independent of speed. Thus, roughlyapproximated weights are sufficient for conceptual calculations andperformance characteristics charts can be produced with reasonableaccuracy. For the large, commercial scale base case design the chartsfound in FIGS. 18 through 20 illustrate expected performance.

The radial wind energy design is a very high performance machine asillustrated in FIG. 18. The base case design at a vehicular speed ofabout 12 meters/sec has the rotating turbines spinning at just overabout 55 RPM with a tip speed of just under about 500 MPH. The chart inFIG. 18 includes a Tip Speed plot that is linearly behaved and where noupper limit constraint has been applied.

The radial design has great flexibility. For example, the generators canbe sized up to maintain enough load on the turbine assembly to limit itsrotational speed to a targeted maximum value. The graph illustrated inFIG. 19 represents a 300 MPH tip speed limitation and extends the plotout to 25 meters per second velocity.

The torque load on the turbines is significantly increased as vehicularspeed increases. This additional torque must be factored into the designspecification of the turbines with or without RPM limitations beingimposed.

As illustrated in the graph of FIG. 20, further reduction of the RPM tolimit tip speed to 200 MPH provides a sense for the difference of powerversus maximum RPM at various vehicle speeds.

Clamping the RPM further to lower the maximum tip speed from 300 MPH to200 MPH reduces the output at a vehicular speed of 25 meters per secondfrom about 72.0 MW down to about 34.3 MW. Thus the importance of adesign capable of highest tip speed is apparent as a one-third reductionin tip speed results in about a 52.4 percent reduction in powerproduced.

If the load the generator is providing is lost for any reason, a rapidrunaway condition exists that may lead to RPM values exceeding designtolerance.

Redundant instrumentation is desired and emergency braking and windstream diversion of all vehicles on the same track should occursimultaneously if any unit exceeds the design maximum RPM value.Maintaining an even following distance between all vehicles whilesuddenly slowing them to a stop may need computer control. Usage ofg-meters as part of the control scheme should be considered along withaccurate and fast responding axel rotation measurement.

Over-the-Road Vehicles

Nothing about the radial designs presented in this patent applicationare conventional and that certainly applies to the mobile designextended to over-the-road vehicles. Both personal transports and trucksare amenable to utilizing the mobile design previously discussed. Thecollector can be scaled down greatly and the speed at whichself-sustained propulsion is achieved becomes greater. For large trucks,the self-sustained speed may never be realized; however, the wind energymay be used to supplement the primary engine and reduce fuel usage.Private vehicles might be equipped with battery capacity greater thanthat needed solely for vehicular needs. The excess capacity can be usedto supplement power needs of the residence provided average drivingfrequency, speed, and distance justify the added cost associated withincreased battery size. A collector design that retracts and reduces itsopening area can be considered so that at highway speeds less throttlingor air bypass is required. Thus reducing the collector's contribution tovehicular drag and reducing the demand on the generator.

Averaged values of weight, cross sectional area, vehicular weight, androlling resistance for small, sports car-type vehicles and formedium-sized cars were combined with approximations for theturbine/collector power output, component weight and drag coefficient togenerate the vehicular performance expectation graphs found in FIGS. 21and 22. For example, a 4-foot diameter collector inlet can be used forboth cases. The turbine assemblies are not exactly scaled down versionsof the “base case.” The simulation runs were for a more compact designwith a larger maximum vane height to minimum vane height ratio. Where aratio of about 1.5-to-1 was used in the “base case,” the automotiveestimation uses a more aggressive about 2.32-to-1 vane height ratio. Ascan be appreciated, other ratios can be used for the base case and theroad vehicle case. As the squash ratio is increased, all post collectorradial dimensions can be reduced. These include the distributor outerdiameter, the turbine diameters, the diffuser diameters, and the shrouddimensions. The bulk fluid model assumes the air expands and contractsvertically as fast the vane profile shape changes. Also, a morefavorable collector bend angle of about 45° (compared to the base case90° angle) was used for the automotive application; however, this is notrequired. A notable size related performance difference between the twocars is the point where self-sustained operation begins. As illustratedby the graphs of FIGS. 21 and 22, the smaller car becomes self-sustainedat approximately 39 MPH and the medium sized car becomes self-sustainedat approximately 43 MPH. FIG. 28 also illustrates how the vehicle'saerodynamic properties can accentuate the collector. Using the hood andwindshield to effectively enhance the collector's function will allowreducing the overall height of the vehicle with collector.

Large tractor-trailer rig transports are also amenable to the benefitsof the radial design. Due to the increased power requirements andspatial limitations, two radial turbines are suggested as theconfiguration of choice; however, this is not required. These units aremounted side-by-side above the cab and utilize the most efficientstraight collector shape. The overall expectation is that theperformance gain of the straight collector and the performance loss ofthe more constrained discharge balance out. As illustrated in the graphof FIG. 23, self-sustained operation is predicted to occur near 78 MPH.The following chart shows estimated performance using a drag coefficientof 0.6 for the tractor-trailer rig. The estimated drag coefficient forthe turbine collector is 0.989.

There are many options for heavy trucks utilizing wind energysupplementation of the primary power source. A hybrid configuration isattractive. Collector angles can be adjusted until the optimal flowperformance for a given sized vehicle is found. The previously discussedinducement of a negative pressure draft condition at the outlet of theturbines is a potentially very rewarding performance enhancement. Thesize and pulling power limitation is directly related to both the totalcollector(s) inlet cross sectional area and the ratio of the sum ofcollector inlet cross sectional area(s) and the vehicle's frontal crosssectional area. Size is important due to wall effects becoming a lesserpercentage of the pressure drop as size is increased. One turbine withcross sectional area “A” will outperform a combined two turbines eachwith “A/2” cross sectional area—while all 3 turbines are equallyoptimized for the flow conditions.

Futuristic Rail System

Due to the diameter needed for a collector to generate enough power topropel a train, a different track width may be utilized. The collectorcan be scaled down from the previously discussed “base case” example inorder to meet both power demands and operational speeds that arepractical for both cargo and personal conveyance. Autonomous vehicles onthe track can be continuously present to provide power to the electricaldistribution grid and the track itself can serve as a component of thegrid. Interspersed amongst the power generation vehicles can be trainsfor delivery of goods and public transportation.

Fan, Blower, and Compressor Applications

Opposite the usage of power extraction from a motive fluid to drive agenerator is the application of power to increase a fluid's velocityhead. Typically, a single motor is needed to drive the turbines;however, very large applications may be configured similar to the windenergy layout where multiple motors can be utilized.

Most fan applications are expected to utilize combined stage mountingswhere all like stages are physically mounted on a common framework androtate at the same rate. Very specialized applications (e.g., highvolumetric flow rates) can implement a design allowing each stage'srotational rate being proportionally controlled via connective gearingto its adjacent stages and the requisite individualized stagesuspension. These specialized applications duplicate the designflexibility observed in the wind energy application; however, the numberof stages used for a fan application is generally dependent upon therequired mass flow rate and pressure differential. Whereas most windenergy applications are generally three stages, designs that producevelocity head can potentially have a much larger number of stages and noupper limit has been established.

The terms ‘bending’ and ‘straightening’ do not apply as precisely indescribing applications that impart energy into the fluid stream. It isconvenient to designate the stage vane shapes by the inlet curvature. Atype “R” vane has a purely radial oriented curvature at its inlet and atangential curvature component at its outlet. A type “T” vane isopposite that of a type R vane. The type T vane has a tangentialcurvature component at its inlet and a purely radial oriented curvatureat its outlet. The type T vanes have the greatest curvature and theirfunction can be considered as “scooping.” The type R vanes are typicallystationary and serve as stationary stream flow conditioning stators;however, in specific applications they can be counter rotating to thestages with which they share a boundary.

The desired exit stream pressure and flow rate dictate the size of theturbine assembly, the number of stages required, the rotational rate,and power requirements. A convenient method of distinguishing betweenclassifying a radial turbine as being a “blower or fan” versus a“compressor” is to consider all single pass radial configurations asbeing either a blower or fan and all multi-pass configurations as beinga compressor. Typically, the flow of the first pass is radially inward(opposite that of power extraction). Most applications will use a type Tstage as the initial stage. Preconditioning the inlet flow bypositioning feed nozzles or baffling in some situations can bebeneficial. To achieve stage-wise pressure increase requires theturbines to be imparting energy into the stream. The relationshipbetween flow rate and discharge pressure is typical of fan curves wherean increased consumer flow demand results in a decreased pressurecomponent of the delivered velocity head. This design is very amendabletowards incorporating a variable speed motor whose usage allows outletpressure control over an expanded flow range as compared to a fixedrotational speed arrangement.

Compressors typically need a greater number of stages than practical ina single-pass configuration. Thus utilizing the radial turbine incompressor applications is accomplished by usage of multiple layers witheach layer having multiple stages. The compressor configuration is thusa layered stack of single-pass radial assemblies. Flow alternatesbetween radially inward and radially outward for each pass. Therelatively high pressure drop required to abruptly reverse flowdirection between each layer is an undesirable characteristic of themulti-pass configuration; however, it is the overall efficiency of thedesign compared against other equipment choices that determines itsviability. The internal flow reversal between layers is achieved simplyby having an open area with floor, ceiling, and center support thatreceives the fluid from one layer and allows it to escape into theadjoining layer. The external flow reversal is accomplished via ahalf-pipe shape bent into a circle. Optional baffling within thehalf-pipe will require optimization and validation through pilotingstudies as the flow characteristics are not easily modeled. The purposeof the baffling is to minimize any tangential flow so that close to pureradial flow is maintained. (Refer to the Internal Combustion Enginedrawing FIG. 25 for an example of a multi-pass compressor.)

Fan and compressor performance is properly defined by measuredexperiment that generates ΔP to RPM performance charts for a range ofcurves each defining a fixed mass flow rate. Based upon wind simulationwork that extracts power, a substantial improvement in efficiency isexpected from the multi-stage radial configurations that are used toimpart power.

Many current applications are expected to benefit by incorporating animproved efficiency primary mover. Of particular interest is thepotential usage in air conditioning applications. If sufficientperformance is realized, water vapor might displace halogenatedrefrigerants as the phase change medium of choice. Water vapor isassigned the R-718 distinction. Water vapor systems are commerciallyavailable and typically require high electricity costs to justify theirmore costly construction expense. Currently, the compressor makes up alarge portion of the increased cost. Very high speed compressors areused in current applications. The radial design is expected to performat greatly reduced rotational speed and the compressor cost is expectedto be significantly lowered.

The illustration in FIG. 24 shows a seven-stage single-pass blower. Fourbending stages rotate as one assembly and the three straightening stagesare stationary.

Note that the impact half of the single cone receiver (functionallysimilar but reversed flow to that of the distributor in power extractionapplications) is part of the rotating assembly. Additional fixedbaffling may be added near the discharge to counter secondary flows. Thecontainment half of the receiver is stationary. These same baffles canbe used to support the impact half of the single cone receiver and thusmake the impact surface stationary.

Internal Combustion Engine

The combination of an inlet compressor, a combustion chamber, and anoutlet energy recovery turbine produces a new type of internalcombustion engine. The compressing, combustion, and energy recoverysteps are analogous to axial flow jet engine operation. The majorincentive for the radial design is its low speed performance advantage.The radial internal combustion engine is expected to operate at greatlyreduced RPM rates and be able to function with a minimum fuel flowrequirement far lower than that of the axial flow jet engine. To achieveenough stages for adequate compression, a multi-layered configurationmay be required. FIG. 25 illustrates a non-limiting multi-layeredconfiguration.

The rotating stages are hung from one common platform that alsoincorporates a gear for transferring power to the generator. Therotating assembly consists of type T stages in the compressor sectionand type B in the energy recovery section. The combustor is centrallylocated. The stationary stages are affixed to the base of the assemblyand consist of type S stages in the compressor section along with type Rand type H stages in the energy recovery section. Note that eachadjacent layer has radial flow in opposite direction and thus the vanesare also reversed as required due to the common direction of rotation.More specifically, the leading tangential edges for the type T stagesfor the first inlet layer are on the outer radius and the leadingtangential edges in the middle layer are located on the inner radius.This reversal of the vane orientation is required for both thepre-combustor compressing and post-combustor energy capture sections.

Compressor Turbine Stage Letter Designations:

T—Stages with inlet curvature that has a tangential component and outletcurvature that is radially oriented. These stages act as “scoops” toforce flow radially towards the combustor.

R—Stages with radially oriented inlet and outlet with a tangentialcomponent of direction. These stages are typically stationary and theycondition the flow so that the next scooping stage has the mosteffective inlet flow orientation.

Energy Capturing Turbine Stage Letter Designations:

B—Bending stages that receive a radial flow and force the stream to takeon a tangential velocity component. These are rotating stages.

S—Straightening stages that receive flow from bending stages. Thesestages are typically stationary and condition flow direction to make thebending stages most efficient.

H—Hybrid stages that complement their adjacent stage functionality.These stages are needed as a result of the multi-layer configuration.They may bend, straighten, or simply act as conduits.

The motor/generator serves a dual purpose. First, it is used toestablish rotation and air flow prior to initiating fuel flow to theburner and starting the combustion process. Second, it converts thecaptured shaft work into electrical power after operational speeds havebeen established.

The addition of distilled water is optional. It can be used to limitmaximum temperatures within the turbine and allow a greater range ofmetallurgical options to be considered.

All radial designs discussed in this patent are negatively impacted bypressure drop. The abrupt flow directional changes encountered in thetransition from one layer to the next is a compromise necessitated bycompactness of design criteria. Stationary equipment with lesser spatialconstraints will tend to have more stages per layer and a reduced numberof layers. For example, one layer for compression and one layer forenergy capture; however, this is not required.

Hydroelectric Applications

The equipment utilized for extracting power from water is analogous tothe wind application except that the dam serves as both the collectorand the shroud. One non-limiting configuration is illustrated in FIG.26. This equipment includes: (1) a reservoir or source of water, (2) afeed pipe with flow control valve, (3) an inlet distributor that feedsthe inner most turbine, (4) radial turbines, (5) an exhaust diffuser,(6) one or more generators with electrical conditioning equipment andcontrols, and (7) the structure to house and tie all the componentstogether.

The low speed performance of the radial design favors low height dams.The generator can be placed above flood stage or located near therotating turbine assembly ring gear. Cavitation problems are notanticipated due to the extremely low pressure drop of the radial design.

Note that the design illustrated in FIG. 26 is driven by head pressurealone. There is no accelerating inlet collector funnel. Thehydroelectric application's single cone distributor's profile determinesif the stream is accelerated. Based upon simulated flow characteristicsobserved in the wind application, the profile with the least pressuredrop is most likely optimal for low head applications. High headapplications may benefit from some acceleration. The bulk flow model isinsufficient for this type of study and advance CFD analysis andpiloting studies need to be part of the design optimization process. Theperforated ring inlet distributor provides even flow distribution.Advanced fluid flow analysis can be used to optimize the perforation'sindividual size and placement. The turbine bearings, ring gear, andgenerator drive gear can optionally be all located above the water line.

The efficiency of the radial design at low head pressure candramatically increase the number of viable sites. The dams would berelatively low height. Even a few feet of head can produce usablequantities of power provided the flow rate is sufficient. Natural andmanmade tidal basins and canal locks are potentially viable; however,streams with continual flow are desirable.

A recommended design feature for the hydroelectric configuration is tohave the perforated distributor ring slightly smaller in diameter thanthe lower half of the single cone distributor; however, this is notrequired. Thus the perforated distributor ring, the upper half of thesingle cone distributor, and the rotating bending vane assembly can belifted or placed as a single unit.

Liquid Pumping Applications

Reversing the hydroelectric configuration effectively creates a liquidpump suitable for low viscosity liquids. Additional head can be had byadding stages and motor power; however, the expected service would below head and high volume as typical of irrigation, flood control, andwater and sewage treatment applications. Larger pumps can use the offsetmotor and ring gear as illustrated with the hydroelectric configuration.The ring gear configuration allows a multiple motor configuration forhigh capacity applications. Smaller scale pumps can use inline motorconfiguration which provides a more compact configuration. Theperforated distributor is unnecessary for pumping service. The outletvalve may be unnecessary if the discharge is above the elevated liquidlevel, or a directional check valve may be needed to prevent reverseflow when the pump is turned off

A combination pumping/generating capability can be considered as a meansof energy storage. Two retention basins can be used and when excesspower is available water is pumped into the elevated basin. Energy isrecovered by allowing flow from high to low with the motors functioningas generators

Waterborne Vessel Propulsion

The usage of a submerged and forward facing straight shaped collectorfor waterborne vessels is analogous to the previously discussed mobileapplications. The power produced by the collector-turbine canpotentially exceed the power required to propel it through the water.The excess power can be used to supplement the primary motive forceengines. Achieving self-sustained operation appears to be feasible.Conventional drive propellers can possibly be replaced with radialdesign pump jet propulsion. All geometric dimensions are sized andoptimized similar to previously discussed applications. Limiting maximumrpm is most likely necessary. A suggested control scheme for limitingthe rpm and controlling the power produced is to raise and lower aleading, prow-shaped, vertically oriented bar in front of the collectorinlet. The prow's action is twofold. It diverts flow around thecollector and acts as an obstruction. Typical values for seawaterdensity and viscosity are used to generate the graph of predictedperformance. The flow simulation has been simplified to reflectnon-compressible hydraulic flow. The performance figures as illustratedin FIG. 27 have 15% subtracted to account for generator, gearing, andparasitic losses. Enhanced performance using draft and non-rectangularturbine passageways has not been included.

To appreciate the magnitude of performance potential, a modern and verylarge container ship engine power and top speed are referenced on thegraph (FIG. 27) for comparison purposes.

Additional Mechanical Detail

This invention aims at defining the flow path, the flexibility ofgeometric design parameters used to optimize each application, andintroducing some of the many applications of which the radial design canbe utilized. The precise mechanical details are not defined. Materialsof construction, mechanical component stress analysis, and precisecomponent thickness profiles are examples of mechanical specificationsthat are absent.

All of the radial applications require the design to incorporatefeatures to facilitate the assembly, disassembly, and maintenance tasks.Of particular concern is the need to maintain accurate turbine assemblyalignment while turbine assemblies are brought together or separated.The size of the radial turbine assembly dictates the practical solutionsneeded to accomplish the alignment task. The huge wind energyapplications set the upper size limit and the lower size limit might bea fan used for applications such as hair dryers or for cooling acomputer. Suggestions based upon size include the following:

1) The smallest designs might utilize threaded guide rods that areremovable after the final mating task is completed. Precisely sized andpositioned bushings in one half of the assembly housing would engage theguide rods secured to the other half and prevent misalignment duringseparation and reassembly.

2) Mid-sized designs where the turbine assembly is large enough to notbe easily handled one-handed require alignment control more robust thansimple guide rods. Here, equally spaced radial lifting lugs can beconsidered. Threaded rods might be used to in conjunction with thelifting lugs. Each rod is rotated in turn only a slight amount as thetwo major assemblies are slowly separated or mated. Some sort ofmechanism for measuring and maintaining alignment is needed to determinewhich rod to turn next and by how much. This measuring method may beprecision levels or an optical device. Perhaps a low power laser aimingdevice secured to a properly aligned mount and kept on a carefullypositioned target located on the other half of the assembly.

3) Very large scale wind applications may require custom designedauxiliary equipment used for assembly and maintenance tasks.Hydraulically-controlled jacks are a possibility. Computer assistance isa possibility where the alignment aim point deviation is used todetermine which adjustments are needed to establish near perfectalignment throughout the entire disassembly or mating process.Precautionary steps can be taken to block or divert wind flows so thatrotational torque does not occur during the disassembly or mating task.Additionally, large scale designs require well distributed liftingforces applied to specifically engineered locations. A lift withoutproper lifting force distribution may damage or even ruin the rotatingassembly by the action of its own weight distorting components beyondacceptable limits.

The designs presented in this invention have, for the most part, flowdirections opposite equipment configurations that are commonly beingused today. The radial design favors inward flow for imparting energyonto the fluid stream. This is opposite squirrel cage type fans and veryhigh speed radial flow gas phase compressors. Also this flow directionis opposite the flow direction of the common centrifugal type of pump.The radial design favors outward flow for capturing energy from streamsin motion. This flow direction particularly contrasts to that of theFrancis type turbine that flows radially inward while accelerating thestream to conclude with a more axial flow transfer of energy at theoutlet. Ideally, as streams become energy depleted, they should transferremaining energy in a radially outward direction using designs thatminimize pressure drop. The radial design allows great flexibility wherestages can rotate independently or have carefully matched geometry thatallows all similar functioning (i.e., flow bending) stages to rotate asa single assembly. Perhaps the most misguided choice of flow orientationfor an application is the usage of axial flow as found in the windenergy market where huge propeller type axial flow designs are beingused to extract energy from wind with an exceptionally low specificspeed. Of course options must exist before the proper choice ofequipment design can be made and the radial design provides thefundamentally correct flow direction for extracting energy from lowspecific speed fluid streams of either gaseous or liquid phase. Equal toaxial flow being the wrong choice for extracting energy from wind is theusage of axial flow in a majority of fan and blower applications. Fansand blowers typically have near stagnant velocity inlet streams and thusthe lowest possible specific speed. The first qualitative graph (FIG. 2)presented in this application clearly conveys why radial flow is theobvious choice for accelerating near stagnant velocity streams.

The radial design configurations may be the best economical choice forproducing commercial quantities of electricity and does so in a muchmore environmentally sound manner than present alternatives. The designis not flawless, as it does generate heat as its emission byproduct;however, as the supply of fossil fuels and air quality concerns continueto mount, the radial design's advantages have it standing alone as theclear winner. No other mechanical energy capturing device is anticipatedto rival the simplicity and fundamentally correct choice of flowmanagement provided by the radial design.

Transforming the transportation systems can only occur if an alternativeexists. The radial designs provide that alternative. Obtrusive windcollectors will by definition never be sleek and certainly lose out in abeauty contest against the curves found on today's sports cars; however,gaining the ability to travel long distance without the burden of fuelcosts is very advantageous. Advanced and autonomous railway systems arethe smart choice for land based transports of both products and people.The shipping industry is also easily recognized as needing atechnological advancement. Today's ships are huge polluters and energyhogs. The radial designs can reduce or eliminate both of these negativeattributes.

It will thus be seen that the objects set forth above, among those madeapparent from the preceding description, are efficiently attained, andsince certain changes may be made in the constructions set forth withoutdeparting from the spirit and scope of the invention, it is intendedthat all matter contained in the above description and shown in theaccompanying drawings shall be interpreted as illustrative and not in alimiting sense. The invention has been described with reference topreferred and alternate embodiments. Modifications and alterations willbecome apparent to those skilled in the art upon reading andunderstanding the detailed discussion of the invention provided herein.This invention is intended to include all such modifications andalterations insofar as they come within the scope of the presentinvention. It is also to be understood that the following claims areintended to cover all of the generic and specific features of theinvention herein described and all statements of the scope of theinvention, which, as a matter of language, might be said to falltherebetween.

What is claimed:
 1. A turbine system comprising a collector, adistributor and one or more multi-vane radial turbine stages, saidturbine stages either stationary or rotatable about a common axis, atleast one of said stages rotatable about said common axis, saidcollector designed to capture and direct a portion of an open and freelyflowing fluid to said distributor, said distributor designed to evenlydistribute and turn the collated stream it receives into a radiallyoutward flow that is oriented into the entrance of a first turbinestage, said fluid contact with a plurality of vanes as said fluid flowsradially outward from said common axis causing said non-stationaryturbine stage to rotate about said common axis.
 2. The turbine system asdefined in claim 1, wherein said collector has a first opening, a secondopening and a body passageway therebetween, said first opening having afirst cross-sectional area and is designed to receive fluid and directsaid received fluid into said body passageway, said second openinghaving a second cross-sectional area and is designed to allow said fluidto exit said body passageway and said collector, said firstcross-sectional area is greater than said second cross-sectional area.3. The turbine system as defined in claim 2, wherein a central flow pathof said collector body may remain in a straight line congruent with thedirection of the feed stream; or may have curvature up to 90°.
 4. Theturbine system as defined in claim 1, wherein said distributor includesa first passageway whose perimeter contains the fluid stream from thecollector, a second passageway that directs the fluid stream into thefirst turbine stage, and an interconnected passageway body that containsthe fluid stream between said passageways.
 5. The distributor as definedin claim 4, further including a diverter valve that diverts away acontrollable fractional amount of the bulk stream to limit the flow ratehandled by the downstream radial stages.
 6. The turbine system asdefined in claim 1, wherein at least one of said stages has an overallcylindrical geometry, said stage having a first opening, a secondopening, and a body passageway therebetween, said opening primarilyoriented towards said common cylindrical axis, said body passageway ofsaid stage being confined by surfaces that comprises a floor and aceiling when said common axis is vertical with a horizon, position insaid body passageway is a plurality of curved vanes extending from saidfloor to said ceiling and which are designed to with said fluid streamto thereby cause stream's direction of flow to be altered.
 7. Theturbine system as defined in claim 6, wherein the vertical vanes thatspan from said floor to said ceiling incorporate a corner roundingprofile, varying vane wall thickness, varying floor to ceiling height,or combinations thereof.
 8. The turbine system as defined in claim 1,wherein at least one of said non-stationary stages includes a suspensionsystem that allows individual freedom of rotation around said commonaxis.
 9. The turbine system as defined in claim 1, including a diffuser,said diffuser having a first opening, a second opening and a bodypassageway therebetween, said first opening having a firstcross-sectional area and is designed to receive fluid flowing radiallyoutward from said outer most stage and to direct such received fluidinto said body passageway, said second opening having a secondcross-sectional area and is designed to allow fluid to exit said bodypassageway and said diffuser, said first cross-sectional area is lessthan said second cross-sectional area.
 10. The turbine system as definedin claim 9, including a shroud, said shroud positioned near saiddiffuser to reduce interference from fluid that did not enter saidcollector with fluid that is exiting said diffuser, said shroudoptionally includes multiple sections around the periphery of saiddiffusor to optimize usage of a negative gauge pressure draft thatresults from said fluid being displaced around an object placed in itsflow path.
 11. The turbine system as defined in claim 1, furtherincludes a pump, a motor or an electric generator, said rotation of saidone or more turbine stages is designed to power said pump, to power saidmotor, or to drive said electric generator to cause electric power to begenerated by said electric generator.
 12. The turbine system as definedin claim 1, whereas said fluid is provided via a closed conduit in placeof said open collector.
 13. The turbine system defined in claim 1,wherein said one or more turbine stages are driven and consume power forthe purpose of creating pumps, fans, blowers, or compressors.
 14. Theturbine system defined in claim 1, wherein increased stage count isachieved by having multiple layers of turbine stages, flow within eachlayer of stages reverses direction from a preceding layer of stages suchas to from radially inward to radially outward or from radially outwardto radially inward flow.
 15. The turbine system defined in claim 12,including a radial turbine compression of air, a combustion chamber, andradial turbine power extraction to create a radial turbine internalcombustion engine.
 16. The turbine system defined in claim 1, whereinsaid collector, said distributor and said one or more multi-vane radialturbine stages are mounted on a moving platform such that a differentialvelocity between said platform and said surrounding fluid provides avelocity gradient needed for said turbine system to extract power. 17.The turbine system as defined in claim 1, further including a movableframe that is designed to allow said collector, said distributor, saidnon-rotating turbine stages, and a shroud to be connected to said frame,said frame including wheels that enable said movable frame to travel ona road or rail system.